Hydraulic machine



I5 She'ets-Sheet l E. K. BENEDEK HYDRAULIC MACHINE Original Filed July 21, 1953 Dec. 7, 1937.

Dec. 7, 1937. E. K. BENEDEK HYDRAULIC MACHINE Original Filed July 24,1933 5 Sheets-Sheet 2 [left 15 fierzedek Dec. 7, 1937. E. K. BENEDEK2,101,732

HYDRAULIC MACHINE Original Filed July 24, 1 953 3 Sheets-Sheet 3grwc/wfdb,

Patented Dec. 7, 1937 UNITED STATES PATENT OFFICE 991. Divided and thisapplication December 3, 1936, Serial No. 114,072

7 Claims.

This application is a division of my copending application 681,991,filed July 24, 1933, now Patent No. 2,074,205, issued March 16, 1937.

This invention relates to fluid pressure generators and moreparticularly to pumps of the kind embodying a plurality of rotatable,radially disposed piston and cylinder assemblies.

The term fluid pressure generator" is herein used for the purpose ofclearness, in order to distinguish the apparatus forming the subjectmatter of this invention from the conventional class of pumps in whichvery high pressure rating is not an absolute prerequisite of theusefulness of the" device. With the imperative demand of natures laws ofeconomy, particularly in regard to the usefulness of a modern machine,it is a primary prerequisite to build the machine so that in a certainsize a maximum amount of power will be obtainable, or the machine besusceptible of a maximum flow of input energy or power. To do this,fluid pressure generators have to operate first at high speed, andsecond at high pressure and thus turn out a maximum possible power fromthe unit. This being an absolute requirement for the usefulness of thistype of generator or hydraulic power machine, I find it convenient touse this new term instead of or descriptive of the term pump, which, ascommonly accepted, now designates a hydraulic machine which sucks anddelivers fluid, practically speaking, at highly difiicult pressures.Therefore, good suction is the main requirement as to the operativenessof such machines. 0n the contrary, a pressure fluid generator isabsolutely useless and inoperative for many present day purposes if itis unable to maintain commercial pressure of the order of 250atmospheres, or the like. For example, a 10 inch diameter ram of ahydro-power press has to exert a force of approximately 1200 tons, andif a pump is not able to create that tonnage and maintain that tonnagefor a determined length of time, it will not be useful and commercialfor that press and therefore will not be used as a generator for thatpress system.

In my co-pending application, Serial No. 681,991, filed July 24, 1933,of which the present application is a division, there is disclosed afluid pressure generator of such construction as to eliminate many ofthe difficulties, limitations upon utility, and other shortcomings ofprior art pumps. Briefly stated, the apparatus disclosed in the parentapplication embodies, among other improvements, a novel arrangement forcoupling the reciprocable members of radial rotatable piston andcylinder assemblies with a rotary reactance member or assembly, sucharrangement being claimed in said parent application; and it alsodiscloses, in common with this application, an improved arrangement formounting and operatively connecting a. pintle and a rotor in a casing,which constitutes the subject matter claimed in this divisionalapplication.

In pumps or pressure generators of the rotatable radial piston andcylinder assembly type the enormous piston pressure up to 500atmospheres, sometimes encountered, creates an equivalent inwardlyacting force, which reacts against a central cylindrical valve known asthe pintle in the American art. This pintle in the pumps of the priorart usually is a long member, supported in the manner of a cantileverbeam. This cantilever beam pintle is.also supported along the rotatingcylinder barrel, being, however, free to flex under the periodicpressure of the resultant piston load, according to the laws of elasticdeflections and vibrations. The elastic vibrations and excessivefriction created between barrel and pintle combines with the similaraction of the piston thrust load transmitting crossheacl pin and createsa characteristic radial pump. noise under pressure, as well known tothose familiar with this art.

To minimize the danger of the hard rubbing mechanical contact betweenpintle and barrel and aforesaid periodic vibrations of the pintlecreated by the periodic piston loads, I provide a novel bearing supportfor the pintle at the end thereof which is opposite to the walled-in oranchored end.

According to the well recognized laws of beam deflection, theunsupported end of a cantilever beam will have a deflection under agiven uniformly distributed load about 67 times as great as the maximumdeflection of the same beam supported at both ends and the same load.

One object of my invention is to provide an apparatus of the classreferred to in which a cantilever pintle is anchored rigidly directly onand immovably with respect to the casing and in which the cantileverpintle nevertheless has a deflection no greater than it would have ifsupported at both ends directly in the casing.

Another object is to provide an apparatus of the kind referred to inwhich a novel arrangement of antifriction bearings is provided formaintaining the pintle and the rotor or cylinder barrel in accurateaxial alignment.

A further object is to provide a novel organization of pintle, rotor andinterposed antifriction rolling bearings so arranged and of such form asto provide a fluid seal to prevent fluid slip between the rotor and thepintle, as well as functioning as bearings to center the rotor withrespect to the pintle.

It will be understood that the foregoing statement of some objects isnot inclusive of all objects, and other objects will become apparentfrom a reading of the following description, the accompanying drawings,and the appended claims. each claim constitutes the attainment of atleast one object of the invention.

In the drawings:

Figure 1 is a longitudinal sectional view of apparatus constructed inaccordance with the invention, the section being taken on the line i--lof Figure 2;

Figure 2 is a section taken on the line 2-2 of Figure 1, with some partsomitted;

Figure 3 is a detail view drawn on an enlarged scale showing a pistonand coupling element assembly partly in elevation and partly in section;

Figure 4 is a top plan view of the assembly shown in Figure .3;

Figure 5 is a perspective view of a piston head. drawn on an enlargedscale;

Figure 6 is a transverse sectional view of a piston actuating rotarymember, the section being taken-on the line 6-6 of Figure '1:

Figure 7 is a longitudinal sectional view taken on the line 1-1 ofFigure 6; and

Figure 8 is a fragmentary vertical sectional view taken on the line 88of Figure 1.

For the sake of simplicity, the illustrations show a hydraulic generatoror pump with an even number of pistons, which, however, may beconstructed with any other number of pistons, as hereinafter will bespecificallydescribed.

Referring to Figures 1 and 2, the pump selected for the illustration ofthe invention comprises a liquid-tight casing comprising a body portion(not shown), a shaft-end cover I, and main covering head 2.

A drive shaft 3 having an annular shoulder 4 and hollow end portion 5 isrigidly secured by a conventional key connection, not shown in thedrawings, to one of the supported ends of cylinder barrel or primarydriving member 6', which end is pulled tightly against shoulder 4 ofdrive shaft 3 by the lock nut 1. Cylinder barrel 8 is supported at bothends on the inner ring of conventional antifriction bearings 8 and 9respectively. The outer races i0 and ii respectively of said bearingsare mounted in covers on heads 2 and l as shown in Figure 2. The pintleI2 is stationarily held through its enlarged portion l3 in the hubportion Id of the main head cover 2 by being pressed into said hubportion i4 and in addition, by means of a conventional key connection,not shown in the drawings.

The central portion of pintle I2 is adapted to distribute the workingfluid and sustain the inwardly acting hydraulic pressure. To provide forthis and minimize all possible friction, deflection and periodicvibrations of the pintle i2, its ends on both sides of its suction andpressure ports I! and i6, respectively, have reduced bearing portions asat l1 and i 8, respectively. Bearing portion i1 cooperates with the endportion IQ of the barrel 6 in the zone of bearing race 8, as a pilotantifriction bearing by means of interposed needle bearings 20. Reducedportion i8 is similarly disposed in the working zone of antifrictionbearing race 9 in such a manner that it projects into the hollow end 5of the drive shaft 3, which forms The provision of the structure definedin the outer bearing race for the needle bearings 2! of this pilotbearing. Thus, it is evident that while the cylinder barrel 6 issupported at its two ends in antifriction bearings 8 and 9 respectively,

the central valve or pintle is also indirectly supported on samebearings through the provision of pilot needle bearings 20 and 2i,respectively, Thus, pintle i2 does not rely on the support'of itsenlarged extension l3, as hitherto has been customary, in the manner ofa cantilever beam, 1. e., on one end only. Pintle i2 has its reversibleports I! and I6 formed in the conventional manner with the exception oftheir contour which provides a continuous curve of a butterfly. Port I6is in communication with port 22 throughaxial channels 23 and 24, whichfurther communicates with main line of the outer circuit. Ports i6 and23 similarly communicate through axial passages 21 and 28, to theoutside oil circuit. Cylinder barrel 6 outside of its wider disc portion23 is provided with a narrow outermost disc portion 30. Each cylinderbore 3i thus extends radially from the outside of the narrow rim 30 tothe inner cylinder port 32 in a radial direction,.to receive the piston33. Each piston hasan opening or eye 34 the parallel flanges 38 and 39of the slide 40.

The slide 40 is provided with ends 4i and 42 which are engaged by theslots 43 provided in the narrow disc portion 30 of cylinder barrel 6.Guide slots 43 are parallel with their respective piston bore so thatslide 40 through its ends 4i and 42 is thus guided radially and driventangentially or angularly by the barrel 6. The crosshead pin 44 isengaged by eye or opening 34 of the piston 33 and has lateralextensions, as at. 45 and 46 to engage the sides 38 and 39 of the slide40 and thereby operate the piston during suction stroke only.

The slide 40 as shown in Figures 3 and 4 is rectangular and comprisesthe slides 38 and 39 interconnected at their ends by the ends 4i and 42,thus forming an open frame. A transverse bore in the sides 33 and 33 isadapted to engage the extensions 45 and 46 of cross pin 44, and therebythe piston 33. Thus, it will be seen that the piston has a free rockingmounting on its crosshead pin 44 or its extensions 45 and 46, therebyensuring free alignment in its respective bore 3i without strain orstress.

An eccentrically mounted piston actuating assembly comprises connectedparts 41 and 48.

It will be seen that when under heavy hydrostatic pintle load, withgreat torque on the slide assembly at full load and full volume, thedriving force between cylinder barrel extension 30 and piston actuatingeccentric parts 41 and 48 tends to rotate or twist the slides about thecrosshead pin 44. In order to accommodate freely for this elastic twistof the slides or slide assembly, the shoulders 36 and 31 of the pistonsshown in Figure 5 are formed arcuately, afiording rocker bearings. Thus,each piston under its maximum load will be perfectly free from thedriving stresses which will be entirely exerted by the slide assemblyitself as coupling member between cylinder barrel 30 and pistonactuating eccentric members 41 and 48, which form a secondary drivenmember or reactance rotor. In order to clearly illustrate how the slides40 are engaged by the eccentric discs 41 and 48, I show in Figures 6 and7 the eccentric itself. According to these figures the assemblycomprises the two eccentric discs 41 and 48, which are bolted togetherby'a plurality of cap screws 49, as shown. Each eccentric disc 41 .or 48is provided with a concentrically disposed circular flange portion as at50 and respectively, ail'ording mountings therefor, as will be set forthlater on. The eccentrio discs have mating. circular faces as at 52 whichare slightly staggered to keep their concentricity the same. Theeccentric assembly or secondary driven member is provided with circularchamber 53 which will accommodate the flange rim 30 of the cylinderbarrel movable therein for the purpose of fluidcontrol. Adjacent and incommunication with chamber 53 are a plurality.

of chordal grooves forming chordal slideways as at 54 and 55 in theeccentric discs, having normally transversely aligned disposition toreceive the longitudinal flanges 38, 39, respectively, of the slidestructures. Each slideway thus has a load transmitting straight bearingsurface 56 and a guide and suction surface 51 to guide the slides duringits operation.

Thus, in Figures 6 and 7 we have an axially detachable rigid eccentricor piston actuating member. It is understood that the load transmittingshoulders inside of the eccentric disc members 41 and 48 may be madeseparably from the disc members and inlayed and fastened inside of saidmembers in a proper manner, or not fastened at all, but kept floatinglyin proper relation by the slides as spacers. Eccentric members 48 and 41are mounted on antifriction bearings 58 and 59 in a well known manner,by having the inner rings 60 and iii mounted directly on ring flanges 62and 63 respectively, and the outer rings,

64 and 65 arranged in bearing retainer rings 68 and 61. Retainer rings66 and 61 are provided with diametrically opposite parallelbearingsliding surfaces as a resultant crosshead, and are supported inmating parallel bearing surfaces of end covers I and 2 respectively asknown in the art. Therefore, they are not shown separately in thedrawings.

Bearing retainers 6B and 61 are shifted by yokes 68 and 69 respectively,they being connected in one rigid assembly by cap screws Ill.

Control rods H and i2 connect the control means to the yoke assembly andcontrol the stroke during operation according to the nature of the jobfor which the pump will be used.

The pump will operate in a well known manner. When the cylinder barreland piston crossheads are in concentric relation, the pistons will standstill in their cylinders and no pumping action will take place. However,when the piston actuating members 41 and M8 are adjusted by the controlrods 'H and 12 to one side or the other of the center of the cylinderbarrel or pintle, the pump will deliver through passages 22 and 26,according to the relative position of the primary and secondary rotors.

The assemblage of parts which has been hereinbefore set forth involvesthe provision, generally speaking, of the primary driving member orpiston barrel 6 associated with the secondary driven member consistingof the eccentric or piston actuating unit including the parts M and 68,the said primary and secondary members or units equipped with the usualpistons 33 carried by the primary driving member 6, and the secondarydriving member or unit having means for the actuation of the pistons inthe manner set forth.

In accordance with the invention, the rolling needle bearings 20 and 2!interposed between the pintle and the cylinder barrel are of elongatedshape and the ratio of their length to their individual diameters, andtheir closely contiguous positioning is such as to provide betweenadjacent rolling elements a plurality of capillary tube spaces.Generally, the desired capillary characteristics will be provided if thetotal length or longitudinal span of the needle rollers is approximatelyequal to the diameter of the pintle. In the apparatus shown as anillustrative embodiment, such characteristics are obtained by usingsingle sets of, needle rollers in unit lengths approximately equal tothe pintle diameter, but it will be understood that the arrangement maybe modified somewhat, depending on the size of related parts,particularly the pintle. In any case,the total bearing extent of needleelements at one end of the pintle should be approximately equal to thepintle diameter. The spaces between contiguous needle rollers willattract and hold bodies of oil, which, in combination with the rollingelements orneedles and the associated pintle and barrel surfaces, act asa fluid seal to, prevent flow of fluid or slip between the pintle andbarrel. Thus, the needle bearings attract fluid by capillary action, soas to maintain a him of oil between the pintle and the barrel,

yet they prevent flow completely along the pintle, i. e., past thebearings. These needle bearings thus have an extremely importantfunction. They have about ten times the load capaciy of any otherbearing that could be supplied and distribute the load on the pintle andbarrel. They maintain a very close spacing of the barrel and pintle sothat the space can be truly capillary and receive slip fluid instantly.They have a large bearing surface themselves and not only attract, butretain oil by attracting and establishing a capillary oil film initiallyaround the pintle. These needle rollers should be hardened and groundand preferably, the total or aggregate circumferential ,spacing betweenthe rollers of each .assembly should be not more than the diameter ofone roller. I have found that this arrangement results in eliminatingsubstantially all eccentricity of the pink: and barrel, with resultantincrease in load capacity.

In some prior art structures, ordinary ball or rolling bearings areinterposed between the ends of the pintle and the barrel. As higherpressures are utilized in such structures, not only is the pintlesubjected to greater hydrostatic unbalance, but also the clearanceitself must be reduced to prevent excessive slip.

In order both to reduce the clearance and yet antifrictionally centerthe pintle in the barrel, a much greater bearing capacity of theantifriction mounting is required. At the same time. at higherpressures, greater heats are developed in the hydraulic fluid and in thepintle and rotor and if any mechanical contact occurred, this heatingeffect would be greatly augmented. In order to solve all of theseproblems concurrently, the specialized capillary needle rollersillustrated are provided between the pintle and barrel.

The advantages of the needle rollers are as follows: First, the use ofneedle rollers permits a tremendous reduction in radial dimension because of the small diameter of the needles. Secondly, due to the factthat ordinary commercial bearings necessarily have relatively greaterradial clearance, whereas needle rollers need only have the veryslightest operating clearance and commercial bearings must allowclearances and tolerances for both the rollers and the races,

whereas the needles require only clearance between them and the pintleand barrel, an additional decrease in radial dimension is permitted.Thirdly, in the prior art structures, only a relatively few rollers canbe provided in a race of given diameter, whereas with the same diameterfrom ten to twenty times as many individual needle antifriction rollerelements can be provided, depending upon the diameter of the latter,which at all events is small. This latter permits not 'only a saving inradial dimension, but increases the actual bearing surface between thepintle and the barrel from ten to twenty times, so that much greaterloads and pressures can be withstood without distortion of the elementsthemselves and without distortion of the cooperating parts between whichthe elements are interposed, and without concentrated stresses andelastic deformation, which latter increases frictional heat..

In addition to the advantages of greater bearing capacity due to thelarge surface of contact, there is also provided a much greater path forthe metal-to-metal conduction of heat from the space between the pintleand barrel; that is, from the slip fluid, so that this heat which wouldotherwise be entrapped may be conducted through the bearings themselvesand more rapidly dissipated. Thus, by using the needle rollers, theradial clearance between the pintle and barrel may.

be reduced. This, in turn, causes a great reduction in slip and thereduction in slip, in turn, causes a great reduction in heat as wellas,an increase in efliciency.

Again, the actual wear or instantaneous deformation of the needlerollers themselves and of the cooperating parts in engagement therewithis eliminated, thus effecting further reductions in the generation ofheat, greater accuracy in the clearance, with a co-existing reduction inthe chances for mechanical friction. Thus, by the mere change fromcommercial rollers to needle roller bearings, much higher efficiency dueto reduced slip, much cooler operation with re-' sultant greateruniformity in operation is obtained. As a result of permitting thecloser clearances and the other features -mentioned, greater pressuresare obtainable and an increase in the pressure of the fluid isaccomplished by an increase in the mechanical eiliciency of thestructure.

In the larger commercial bearings, regardless of whether caged or not,the entire load must be carried on a very comparatively few spacedpoints of contact. Under the heavy loads and pressures involved inpressure fluid generators and motors, the concentration of load due tothe widely spaced points of contact results in extreme elasticdeformation, both of the antlfriction elements and the cooperatingparts, not only developing heat, but permitting misalignment of otherparts of the apparatus, particularly the barrel and pintle where theclearance should be uniform and as small as possible without danger ofseizure. In connection with the pintle and barrel clearance, it shouldbe noted that with absolute concentricity of the barrel and pintle witha resultant uniform clearance circumferentially of the pintle, the slipmay have a fixed value based on the pressure :69!) for each recharge.

In such structures, also, there is a periodic vibration as each pistondischarges or receives its charge. Thus, with five pistons at 1200 R. P.M., there is a periodic vibration of a frequency of 600 per minute foreach discharge of the piston and i This, in effect, creates a frequencyof 20 vibrations per second. These vibrations are reflected in the spacebetween the pintle and barrel with the result that they must bewithstood directly by the needle rollers. Consequently, the bearings 20and 2i must not only rotate at high speed, but concurrently must becapable of resisting these periodic shock forces of the hydrostaticpressure. 7

In addition to these more apparent advantages of the needle rollermountings of the pintle and barrel, attention is directed to thefollowing operating effects: the needle rollers at the ends of thepintle resist shock loads of the pressure cycles, thereby moreaccurately maintaining the clearance space between the pintle andbarrel. By maintaining this clearance space between the pintle andbarrel, they permit the limited capillary spacing of the pintle andbarrel so that a capillary fluid film forms and provides hydrostaticbalance on the opposite sides of the pintle. Further, they provide apressure seal in two directions between the pintle and barrel during thepressure cycle. This results from the fact that they maintain acapillary oil film which is highly tenacious.. Further, they effect asuction seal in both directions during the suction cycles, thuseliminating the entrance of air between the barrel and pintle andsubsequently therefrom into the fluid circuit.

Due to their capillarity, they maintain lubrication between the pintleand barrel not only at the bearings themselves, but elsewhere along theentire valve portion during idle or zero stroke periods of the pump ormotor, as well as during operation. Due to their capillary action, theyprovide a fluid cushion for purposes of starting when the pump isstanding idle and effect prelubrication to prevent instantaneous seizureof the pintle and barrel upon starting and scoring of the pintle bycontact with the barrel or particles of foreign matter accumulatedbetween the pintle and barrel.

This latter results from maintaining lubricant between the barrel andpintle at all times. Due to the fact that an oil seal can be obtained bythe bearings without completely blocking the end of the barrel bore,foreign particles can be discharged gradually over a period of time fromthe barrel bore. Due to the effective oil seal provided by the bearings,not only is the operating oil film between the valve portion of thepintle and barrel bore maintained during idle periods and prevented fromdraining away but also the gradual drainage of fluid from an individualcylinder during stop periods is prevented, so that air does not enterand replace the fluid. In case of ordinary bearings, such fluid wouldquickly drain from the cylinders and escape so that upon starting theoperation, the motor would be subject to air entrapment and wouldrequire a. considerable period of running until the air was entirelydischarged from the system.

The seal provided by the bearings is efiective during operation bothagainst excessive slip or air which might otherwise be occasioned bymachining inaccuracies or wear, thus greatly extending the useful lifeof the pump or motor. In view of these differences, it is readilyapparent that the interposition of the capillary needle rollers betweenthe pintle and barrel perform, in this new surrounding, a functionseparate and apart from their function as bearings, which functionresults from a new cooperative relation with the barrel and pintle, asdescribed above. For example, the effective seal against the escape ofoil or the intake of air in the clearance space between the pintle andbarrel is a result effected only in connection with the particularstructure with which the bearings are associated. In fact, they makepossible the elimination of the usual packing gland used for sealingbetween the barrel and pintle for retaining fluid, which glands areusually provided in the ordinary types of bearings.

It will'be understood that various changes may be made in theconstruction and relative arrangement of the parts without departingfrom the invention as defined in the claims.

I claim:

1. In a radial piston pump or motor, a rotor, piston and cylinderassemblies carried thereby, means to actuate the assemblies uponrotation of the rotor, said rotor having an axial bore, a valve pintlereceived in said bore and having ports for cooperation with theassemblies, and elongated cageless capillary needle rollers interposedbetween the barrel and pintle, and said rollers being freely rotatablewith respect to each other and anti-frictionally constraining the pintleand barrel to coaxial relation.

2. In a radial piston pump or motor, a rotor, piston and cylinderassemblies carried thereby, means to actuate the assemblies uponrotation 06: the rotor, said rotor having an axial bore, a valve pintlereceived in said bore and having a valve portion with ports intermediateits ends for cooperation with the assemblies, and elongated cagelesscapillary needle rollers interposed between the barrel and pintle'ateach end of the valve portion, and said rollers being freely rotatablewith respect to each other and anti-frictionally constraining the pintleand barrel to coaxial relation.

3. In a radial piston pump or motor, a rotor having an axial bore; apiston and cylinder assembly carried by said rotor; means to actuatesaid assembly upon rotation of said rotor; a valve pintle received insaid bore and having a port for cooperation with said assembly; andmeans for maintaining said rotor centered with respect to said pintleand for obstructing the slip of fluid between the pintle and the wall ofthe rotor bore beyond one end of the pintle comprising a set of closelycontiguous elongated anti-friction bearing rollers interposed betweensaid 'pintle and the wall of said bore adjacent one end of said pintle,the ratio of the length of said rollers to their individual diameter andthe close spacing of said rollers being such as to provide between therollers a plurality of capillary tube spaces adapted to maintain thereinbodies of oil by capillary attraction to thereby provide a combinedanti-friction bearing and fluid seal.

4. In a radial piston pump or motor, a rotor having an-axial bore; apiston and cylinder assembly carried by said rotor; means to actuatesaid assembly upon rotation of said rotor; a valve pintle received insaid bore and having a port intermediate its ends for cooperation withsaid assembly; and means for maintaining said rotor centered withrespect to said pintle and for obstructing the slip of fluid between thepintle and the Wall of the rotor bore beyond the ends of the pintlecomprising at each end of the pintle and respectively on opposite sidesof said port a set of closely contiguous elongated anti-friction bearingrollers interposed between said pintle and the wall of said bore theratio of the length of said rollers to their individual diameter and theclose spacing of said rollers being such as to provide between therollers a plurality of capil- Y lary tube spaces adapted to maintaintherein bodies of oil by papillary attraction to thereby provide acombined anti-friction bearing and fluid seal.

5. A construction as set forth in claim 3 in which at least one of thefollowing elements therein: namely, the rotor and the pintle, is formedwith a recess for receiving said bearing rollers, the depth of therecess being no greater than the individual diameters of the bearingrollers.

6. A construction as set forth in claim 3 in which the rotor and thepintle are formed with radially aligned recesses for receiving saidbearing'rollers, the combined depths of said recesses belng no greaterthan the individual diameters of the bearing rollers.

7. In a radial piston pump or motor, a rotor having an axial bore; apiston and cylinder assembly carried by said rotor; means to actuatesaid assembly upon rotation of said rotor; a valve pintle received insaid bore with sufiicient clearance to provide an intervening space foraccommodating a body of oil, said pintle having a port for cooperatingwith said'assembly; and means for centering said rotor with respect tosaid pintle and for maintaining a body of oil in said intervening spacewhile preventing flow or slip of oil therethrough comprising a set ofclosely contiguous elongated antifriction bearing rollers interposedbetween and operatively engaging the pintle and the wall of said bore,the ratio of the length of said rollers to their individual diametersand the close spacing of said rollers being such as to provide betweenthe rollers a plurality of capillary tube spaces whose capillaryattraction draws oil from the adjacent portion of the space between thepintle and the rotor but resists flow or slip of oil completely pastsaid bearing rollers.

ELEK K. BENEDEK.

